In a hydraulic power steering apparatus which assists the steering operation by hydraulic force generated by a plurality of hydraulic cylinders (power cylinders) disposed in the steering apparatus for reducing labor required for the operation of a steering wheel, a hydraulic control valve for controlling supply and discharge of oil pressure in accordance with direction and magnitude of a steering torque applied to the steering wheel is disposed between, a hydraulic pump driven by an electric motor and a hydraulic tank for accommodating operating oil and the hydraulic cylinders, and pressurized oil generated by the hydraulic pump is supplied to two cylinder chambers of the hydraulic cylinders by the operation of the hydraulic control valve.
As for the hydraulic control valve, a rotary-type hydraulic control valve directly utilizing the rotation of the steering wheel is used. In this rotary hydraulic control valve, an input shaft connected to the steering wheel and an output shaft connected to the steering mechanism are coaxially connected to each other through a torsion bar, a valve spool integrally formed with one of the connected ends is coaxially and relatively rotatably fitted into a cylindrical valve body engaged with the other connected end, and when the steering torque is applied to the steering wheel, the torsion bar is twisted and relative angular displacement is generated between the valve body and the valve spool.
FIG. 1 is a schematic transverse sectional view showing one example of a structure of a conventional hydraulic control valve described in Japanese Patent Application Laid-open No. H9-39814 (1997), and FIG. 2 is a schematic transverse sectional view showing another example of the structure of the conventional hydraulic control valve described in Japanese Patent Application Laid-open No. H9-39814 (1997).
In FIGS. 1 and 2, a plurality of first oil grooves 4 extending in the longitudinal direction are disposed along the circumferential direction in an inner peripheral surface of a valve body 1. A plurality of second oil grooves 5, 5, . . . are alternatively disposed in an outer peripheral surface of a valve spool 2 with respect to the first oil grooves 4, 4, . . . .
In the example shown in FIG. 1, the first oil grooves 4, 4, . . . parallelly arranged in the inner peripheral surface of the valve body 1 form first oil feed chambers 12, 12, . . . and second oil feed chambers 13, 13, . . . alternately. The first oil feed chambers 12, 12, . . . are in communication of a right side cylinder chamber SR of a hydraulic cylinder S to which oil is fed through an oil feed hole formed in the valve body 1. The second oil feed chambers 13, 13, . . . are in communication with a left side cylinder chamber SL of the hydraulic cylinder S through an oil feed hole formed in the valve body 1. Further, the second oil grooves 5, 5, . . . parallelly arranged in the outer peripheral surface of the valve spool 2 alternately form oil supply chambers 10, 10, . . . which are in communication with discharge side of a hydraulic pump P as an oil pressure source through an oil introducing hole formed in the valve body 1 and oil discharge chambers 11, 11, . . . which are in communication with an oil tank T to receive the oil through an oil discharge hole formed in the valve spool 2.
In an example shown in FIG. 2, second oil grooves 5, 5, . . . parallelly arranged in an outer peripheral surface of a valve spool 2 alternately form first oil feed chambers 12, 12, . . . and second oil feed chambers 13, 13, . . . . The first oil feed chambers 12, 12, . . . are in communication a right side cylinder chamber SR of a hydraulic cylinder S to which oil is fed through an oil feed hole formed in the valve body 1. The second oil feed chambers 13, 13, . . . are in communication with a left side cylinder chamber SL of the hydraulic cylinder S through an oil feed hole formed in the valve body 1. First oil grooves 4, 4, . . . parallelly arranged in an inner peripheral surface of the valve body 1 alternately form oil supply chambers 10, 10, . . . which are in communication with discharge side of a hydraulic pump P through an oil introducing hole formed in the valve body 1 and oil discharge chambers 11, 11, . . . which are in communication with an oil tank T through an oil discharge hole formed in the valve spool 2.
In any of these examples, the first oil grooves 4, 4, . . . and the second oil grooves 5, 5, . . . are in communication to each other while having equal sectional areas in the circumferential direction at both sides in widthwise direction, in a neutral state in which the above-described relative angular displacement is not generated, and the communicating portions act as throttle portions 6a, 6b whose throttle areas are changed in accordance with the relative angular displacement. Therefore, oil pressure to be supplied to the cylinder chambers SR, SL through the first and second oil feed chambers 12, 13 are controlled by the change in throttle areas of these throttle portions 6a, 6b. 
Next, the operation of the hydraulic control valve associated with the relative angular displacement.
FIGS. 3A and 3B are explanatory views of operation of an oil feed chamber, an oil supply chamber, and an oil discharge chamber shown in straightly developed manner, which are arranged along a fitted peripheral surfaces of a valve body and a valve spool in the conventional hydraulic control valve.
FIG. 3A shows a state in which the relative angular displacement is not generated between the valve body 1 and the valve spool 2. In this state, the pressurized oil supplied from the hydraulic pump P to the oil supply chambers 10 flows such that the oil is equally divided into the first and second oil feed chamber 12, 13 which are adjacent to each other through the throttle portions 6a, 6a having the equal sectional areas in the circumferential direction on the both sides of the oil supply chambers 10 and are introduced into the oil discharge chambers 11, 11 through the throttle portions 6b, 6b having the equal sectional areas in the circumferential direction on the both sides of the first and second oil feed chamber 12, 13, and flows into the oil tank T which is in communication with these oil discharge chambers 11, 11. Therefore, the oil pressure supplied to the oil supply chambers 10 is not fed to any of the cylinder chambers SR, SL, and no force is generated in the hydraulic cylinder S.
FIG. 3B shows a state in which the steering torque is applied to the steering wheel, and the relative angular displacement is generated between the valve body 1 and the valve spool 2. In this state, the throttle area of one of the throttle portions 6a, 6a on the both sides of the oil supply chamber 10 (the first oil feed chamber 12 side) is increased, and the throttle area of the other throttle portion 6a (the second oil feed chamber 13 side) is reduced. As a result, the pressurized oil supplied to the oil supply chamber 10 is introduced mainly to the first oil feed chamber 12 through the throttle portion 6a whose throttle area is increased. That is, a pressure difference is generated between the first oil feed chamber 12 and the second oil feed chamber 13, i.e., between the cylinder chambers SR, SL which are in communication with the first oil feed chamber 12 and the second oil feed chamber 13, and the hydraulic cylinder S generates hydraulic force (steering assist force) corresponding to this pressure difference.
The pressure difference generated at that time depends on a degree of reduction in the throttle area of the other throttle portion 6a (the second oil feed chamber 13 side), and the degree of the reduction corresponds to magnitude of the relative angular displacement, i.e., the magnitude of the steering torque applied to the steering wheel. Therefore, the force generated by the hydraulic cylinder S has a direction and magnitude corresponding to the steering torque, and can assist the steering operation. At that time, oil in the left cylinder chamber SL pushed out by the operation of the hydraulic cylinder S is circulated into the second oil feed chamber 13, and is introduced into the oil discharge chamber 11 through the throttle portion 6b whose throttle area is increased on the one side of the second oil feed chamber 13, and is discharged into the oil tank T which is in communication with the oil discharge chamber 11.
Meanwhile, preferable increasing characteristics of the steering assist force in the power steering apparatus are not characteristics to increase in proportion to the steering torque, but are characteristics in which the steering assist force is gradually increased in a range having small steering torque but the force is rapidly increased when the steering torque exceeds a predetermined limit. To obtain such characteristics, chamfer portions 7, 7, . . . each having a predetermined inclination angle with respect to a peripheral surface of the valve spool 2 and having a predetermined width in the circumferential direction are formed on corner portions of the first oil grooves 4, 4, . . . side of all of the second oil grooves 5, 5, . . . of the valve spool 2. With these chamfer portions, the throttle areas of the throttle portions 6a, 6b are slowly changed with respect to the relative angular displacement between the valve body 1 and the valve spool 2.
In the hydraulic control valve operating as described above, four, six, eight or more first and second oil grooves 4, 5 are equally disposed, half of one of the first and second oil grooves 4 and 5 are formed as the oil supply chambers 10, and the remaining half are formed as the oil discharge chambers 11.
In a hydraulic control valve in which four oil grooves are disposed at equal distances from one another, since the oil supply chambers 10 and oil discharge chambers 11 are provided two each, the flow rate of the pressurized oil introduced into one oil supply chamber 10 can be relatively large. On the other hand, since the two oil supply chambers 10 whose oil pressure become high because they control the hydraulic cylinder S are disposed with phase difference of 180 degrees, balance of pressure distribution applied to the valve body 1 is inferior, and the valve body 1 is deformed into elliptic shape. At that time, a fitting gap of about 10 μm between the valve body 1 and the valve spool 2 is changed, and there is an adverse possibility that biting phenomenon is generated between the valve body 1 and the valve spool 2.
In a hydraulic control valve in which six oil grooves are disposed at equal distances from one another, the oil supply chamber 10 and oil discharge chamber 11 are provided three each, and in a hydraulic control valve in which eight oil grooves are disposed at equal distances from one another, the oil supply chamber 10 and oil discharge chamber 11 are provided four each. Therefore, the flow rate of the pressurized oil distributed to one oil supply chamber 10 is smaller than that of the hydraulic control valve in which four oil grooves are disposed at equal distances. However, in the hydraulic control valve in which six oil grooves are disposed at equal distances, the oil supply chambers 10 whose oil pressure become high are disposed at three locations with phase difference of 120 degrees, and in the hydraulic control valve in which eight oil grooves are disposed at equal distances, the oil supply chambers 10 whose oil pressure become high are disposed at four locations with phase difference of 90 degrees. Therefore, the balance of the pressure distribution applied to the valve body 1 is excellent, the deformation of the valve body 1 is restrained, and the fitting gap between the valve body 1 and the valve spool 2 is excellently held. That is, it is preferable to dispose at least six oil grooves in the hydraulic control valve.
The present applicant is developing a power steering apparatus in which the conventional hydraulic control valve of the above-described structure is used, a hydraulic pump is standby controlled (low-speed rotation or zero-speed rotation), and when a steering torque is not applied to a steering wheel at the time of idling or the like, the small or zero flow rate of pressurized oil about 1 to 2 liter/min is introduced into the oil supply chamber of the hydraulic control valve, the steering angle of the steering wheel is detected, and the flow rate of oil of the hydraulic pump can be increased in accordance with the steering angular velocity based on the detected steering angle. According to such a power steering apparatus, it is possible to abruptly change the flow rate to be controlled of the hydraulic control valve from the small flow rate as small as possible or zero flow rate to high flow rate as compared with the conventional small flow rate.
FIG. 4 is a view showing the flow rate characteristics of the conventional hydraulic control valve showing a relation between steering angular velocity and the flow rate of a pump, and FIG. 5 is a view showing hydraulic characteristics of the conventional hydraulic control valve showing a relation between steering torque applied to the steering wheel and hydraulic force controlled by the hydraulic control valve.
In FIG. 4, the number of revolution of an electric motor for the hydraulic pump is increased as the steering angular velocity is increased, and the flow rate of the hydraulic pump is straightly increased. In FIG. 5, the increase in the steering torque is reduced as the hydraulic force controlled by the hydraulic control valve is increased.
However, the conventional hydraulic control valve is designed or produced such that the flow rate (minimum flow rate to be controlled) at the time of standby control becomes the small flow rate of about 1 to 2 liter/min or zero flow rate as described above. Therefore, when the conventional hydraulic control valve is applied to the hydraulic control valve under development, the oil pressure characteristics in a region where the flow rate to be controlled is minimum which is smaller than the conventional small flow rate become unstable. That is, when the oil pressure rises at the beginning of steering operation, there are problems that the hydraulic characteristics becomes abruptly discontinuous, and the steering torque becomes discontinuous.
FIG. 6 is a view for explaining a state where the hydraulic characteristics of the conventional hydraulic control valve become discontinuous.
Here, it is supposed that the conventional hydraulic control valve whose hydraulic characteristics in the small flow rate region is varied largely and which has six or eight oil grooves disposed at equal distances from one another is applied to the hydraulic control valve under development as it is. That is, in the case of the hydraulic control valve having eight oil grooves disposed at equal distances from one another, the flow rate to be distributed to each of the throttle portions is reduced as small as 0.125 liter/min (1 liter/min is equally divided into eight) or less, and this small flow rate causes large variation in the hydraulic characteristics, and this is controlled by the chamfer portion and therefore, the rise of the oil pressure characteristics at the beginning of the steering operation becomes unstable. Therefore, when the steering operation is started, “jump” is generated in the oil pressure characteristics as shown in FIG. 6, the hydraulic characteristics become discontinuous, and the steering torque becomes discontinuous.
On the other hand, in the case of the hydraulic control valve having four oil grooves disposed at equal distances from one another, the flow rate to be distributed to each of the throttle portions is increased as compared to a case in which a hydraulic control valve having six or more oil grooves disposed at equal distances from one another, but the balance of pressure distribution applied to the valve body is inferior, the biting phenomenon is generated between the valve body and the valve spool and thus, it is not preferable to apply the hydraulic control valve having the four oil grooves disposed at equal distances from one another as already described above.
The present invention has been accomplished to solve the above-described problems, and it is one object of the invention to provide a hydraulic control valve in which only ones of the first and the second oil grooves facing the throttle portions between the oil supply chambers and the oil feed chambers or facing the throttle portions between the oil discharge chamber and the oil feed chambers are provided with the chamfer portions formed on those corner portions of the ones of the first and the second oil grooves which are closer to the other ones. Therefore, even if six or more oil grooves are disposed at equal distances from one another, it is possible to stabilize the hydraulic characteristics when the minimum flow rate to be controlled is reduced as small as possible, and to eliminate the discontinuity of the hydraulic characteristics in a region where the flow rate to be controlled is minimum.
It is another object of the present invention to provide a power steering apparatus which includes a hydraulic control valve having stable hydraulic characteristics in a region where the flow rate to be controlled is minimum, and the hydraulic pump is driven such that a flow rate becomes low or zero flow rate during the standby control when steering operation is not carried out, and such that the flow rate becomes high flow rate in accordance with steering angular velocity when steering operation is carried out. With this design, it is possible to reduce the energy consumption while the steering wheel is not operated at the time of idling for example without the discontinuity of the hydraulic characteristics.